Mounting for a turbo-machine rotor and its use

ABSTRACT

The invention relates to a mounting for a turbo-machine rotor. Said mounting is characterized in that radial support of the rotor is assured by one or more permanent magnet bearings. One or more single-thrust sliding bearings provide axial support. The permanent magnet bearings also serve as a lifting device for the sliding surfaces of the hydrodynamic sliding bearings. A force, acting in the opposite direction from the axial force, is generated by axial displacement of the rotor-sided bearing elements of the permanent magnet bearing, in relation to the housing-sided bearing elements, counter to the direction of axial force. This force separates the sliding surfaces of the hydrodynamic sliding bearings from each other in the stopping phases of the rotor. Once it has come to a stop, and when it is in the starting and stopping phases, the rotor is guided by one or more sliding or rolling bearings. These are so configured that, at low rotational speed and when at a standstill, they have a substantially lower moment of friction than the hydrodynamic sliding bearings used in the mounting.

The invention concerns a bearing arrangement for a rotor of a fluid flowmachine having at least one permanent magnet bearing for radial guidanceand at least one hydrodynamic plain bearing—which carries the axialthrust of the fluid flow machine—for axial—or for axial andradial—guidance in the operating phase. The invention also concerns ause of the fluid flow machine.

It is known for the lubrication of rolling and plain bearing in fluidflow machines or turbines to be effected in many cases by the flowmedium itself. The service life of the bearings lubricated in that waycan be increased when using wear-resistant bearing materials such as forexample silicon nitride in rolling bearings or silicon carbide in plainbearings.

The use of hydrodynamic and hydrostatic plain bearings results in abearing arrangement without contact of the sliding surfaces in theoperating phase and thus affords a bearing arrangement with a low degreeof wear. Contact-free bearing arrangements are also represented by theuse of electrical or magnetic forces—for example in active magneticbearings in accordance with WO 95/13477—in bearings with superconductormaterials or in permanent magnet bearings. In that respect, no lubricantis necessary to achieve the support action. Permanent magnet bearingsare always used in conjunction with other kinds of bearings as, inaccordance with Earnshaw's law, it is not possible for a static body tobe stably supported in all three directions in space exclusively bymeans of permanent magnets. In most cases they are used in combinationwith an active magnetic bearing. Moreover in regard to active magneticbearings there are constructions in which the rotor is radially guidedby rolling bearings in the starting and stopping or run-down phases inorder to suppress rotor oscillations.

Hydraulic compensating devices such as for example compensating plateswhich with the flow medium produce an axial thrust compensation effectare also known. To promote the axial thrust compensation effect intransient operating phases, EP 0 355 796 describes the combination ofaxial thrust compensation with an active magnetic arrangement. Acombination of a device for axial thrust compensation with a liftingdevice which, by means of permanent magnets, prevents contact of theload-relief plates in the stopped condition, is to be found in EP 0 694696. The permanent magnets in that case do not have a bearing function.

The known bearing structures suffer from a number of disadvantages.Rolling and plain bearings are heavily loaded when involving lubricationby means of media which have a poor lubricating action—for example whichare of very low viscosity—and they attain only short operating lives oroften have to be maintained or replaced. Other bearings suffer from thedisadvantage that additional technical complication and expenditure isrequired for satisfactory operation of the bearings upon starting up andin part also in operation thereof: hydrostatic plain bearings require apressure source which—for example when the bearings are used inpumps—can admittedly in the operating phase also be represented by thepump itself, but which must be present as an external component for thestarting phase. A similar consideration applies in regard tohydrodynamic plain bearings whose high starting moment often has to becompensated by suitable technical measures, for example hydrostaticstarting. Hydraulic compensating devices in previous designconfigurations also require technical arrangements, for exampleadditional bearings, in order to compensate for the axial thrust forceof the fluid flow machine in the transient operating phases and in orderto prevent high starting moments, for example due to compensating platesbeing in contact. Active magnetic bearings require electrical power forthem to operate and they need active regulation as well as additionalauxiliary bearings to cover the case of failure of the magneticbearings.

When using superconductor magnetic bearings, for example when usingcryogenic—intensively cold—flow media, additional complication andexpenditure is necessary in order to ensure the correct position of therotor at temperatures above the superconducting phase of thesuperconductors. When using cryogenic media, with many of theabove-mentioned bearings there is the risk of the rotor jamming due toimpurities contained in the cryogenic medium becoming frozen on, whilethe machine is in the stopped condition.

In consideration of those factors the inventor set himself the aim ofproviding a low-wear bearing arrangement for rotors of fluid flowmachines or turbines, which do not have a lubricant apart from the flowmedium, wherein the rotor bearing arrangement is to be distinguished interms of implementation thereof by a low starting torque and a low levelof technical complication and expenditure.

That object is attained by the teaching of the independent claim; theappendant claims set forth desirable developments. The scope of theinvention also embraces all combinations consisting of at least two ofthe features disclosed in the description, the drawing and/or theclaims.

In accordance with the invention, in the operative position of the rotorthe components of one or more permanent magnet bearing or bearings onthe rotor are displaced with respect to the associated components on thehousing side axially in opposite relationship to the direction of theaxial thrust of the fluid flow machine out of the position of forceequilibrium, and there is produced a force which is directed in oppositerelationship to the axial thrust and which provides that the slidingsurfaces of the hydrodynamic bearings which are operative in theoperative phase are separated from each other during the starting phaseand the rotor is pressed against one or more additional plain or rollingbearings; the latter provide for guidance of the rotor during thestarting and stopping phases and are structurally such that, at lowspeeds of rotation and in the stopped condition, they have asubstantially lower moment of friction than the hydrodynamic plainbearings used in the bearing arrangement.

In accordance with the invention therefore the rotor is guided radiallyby means of one or more permanent magnet bearings. Axial guidance iseffected by one or more hydrodynamic plain bearings which act on oneside, wherein the rotor of the fluid flow machine is pressed against thebearing or bearings by the axial thrust which occurs in the operatingphase. If the hydrodynamic bearings are of a suitable configuration, itis also possible for radial forces to be carried by the hydrodynamicplain bearings in operation of the machine. The permanent magnetbearings serve at the same time as a lifting device for the slidingsurfaces of the hydrodynamic plain bearings.

Axial displacement of the bearing elements on the rotor side relative tothe bearing elements on the housing side produces a force which isdirected in opposite relationship to the axial thrust and whichseparates the sliding surfaces of the hydrodynamic bearings from eachother in the run-down or stopping phases. In the stopped condition andduring the starting and stopping phases the rotor is guided by one ormore plain or rolling bearings which, at low speeds of rotation and inthe stopped condition, have a substantially lower frictional moment thanthe hydrodynamic bearings used; that minimises the starting torque ofthe flow machine.

For the purpose of achieving axial displacement of the rotor into theoperative position in the starting phase with a low level of technicalcomplication and expenditure, the above-mentioned axial displacement ofthe rotor is preferably so limited that the magnetic forces of thepermanent magnet bearings which occur remain lower in the axialdirection than the minimum axial thrust of the machine in operationthereof. As a result, when starting up, the rotor automatically movesinto the operative position.

If limitation of the axial displacement of the rotor is not possible oris not wanted, the rotor or the auxiliary bearing, in accordance withthe invention, can be displaced during the starting phases by means ofan additional device—for example an electromagnet—axially in thedirection of the hydrodynamic bearing or bearings until the axial thrustof the fluid flow machine exceeds the axial magnetic forces of thepermanent magnet bearings and the rotor automatically moves into theoperative position.

To support a rotor of a fluid flow machine with a gaseous flow medium ora flow medium of very low viscosity, the invention also proposes thatthe hydrodynamic plain bearings may be in the form of aerodynamic plainbearings.

To simplify the structure involved and to reduce the gap losses, thehydrodynamic plain bearings in accordance with another feature of theinvention can be connected to cover plates or disks of closed impellers.The lubrication gaps then serve at the same time as sealing gaps for theimpellers. In that respect, it has proven desirable for the slidingsurface, on the rotor side, of at least one hydrodynamic plain bearingto be integrated into the face of an impeller with a cover plate ordisk, and for the plain bearing at that location to be caused to form aseal in respect of the high-pressure side of the impeller in relation tothe low-pressure side.

In accordance with a further feature of the invention the rotor can besupported radially by the permanent magnet bearings and there can beprovided a hydrodynamic axial or thrust bearing near the center ofgravity of the rotor; an auxiliary bearing—in particular a cone-typeplain bearing may be arranged near the center of gravity of the rotor.Advantageously, the auxiliary bearing can be integrated into the rear ofan impeller.

It is also in accordance with the invention for the rotor of an electricmotor driving the fluid flow machine—for example an electric motor inthe form of a synchronous motor with permanent magnet excitation—to beintegrated into the rotor of the fluid flow machine.

For reasons of longer service lives, a further feature of the inventionprovides that bearing or magnet rings of the permanent magnet bearings,possibly also an excitation magnet or a magnet ring of the electricmotor, are protected by preferably electrically conducting protectiverings. At particularly high speeds of rotation, the magnet rings on therotor side should be protected with particularly strong, low-weightbanding or retaining rings—for example of carbon fiber-reinforcedplastic material.

It has also proven to be advantageous to associate at least onehydrostatic plain bearing or a hydraulic compensating device with thehydrodynamic plain bearing in assisting relationship therewith or toreplace the hydrodynamic plain bearing by a hydrostatic plain bearing ora hydraulic compensating device. In accordance with the invention afront-end or intake impeller can also be provided in a magnetic bearing.The intake impeller is then fixed within the bearing ring on the rotorside.

The advantages achieved with the invention lie in particular in the lowlevel of technical complication and expenditure with which a low-wearbearing arrangement which is distinguished by a long service life can beimplemented. There is no need for a feed of electrical or hydraulicpower and there is also no need for a regulating device for operation ofthe bearing arrangement. The bearing arrangement can be produced withfew components, using simple manufacturing technology, and being of lowweight. When an individual aerodynamic bearing is used for axiallysupporting the rotor, there is no need for an additional resilientsupporting arrangement of a sliding surface or elements thereof.

The bearing arrangement permits small gap clearances and thereby makesit possible to minimise gap losses, in particular in the case of fluidflow machines involving flow media of very low viscosity. In operationof a fluid flow machine with cryogenic media, the risk of impuritiesbecoming frozen on in the stopped condition is minimised by the rotorbearing arrangement according to the invention because the pump gaps areenlarged in the stopped condition in comparison with the operativecondition and contact between the rotor and the housing occurs only inthe region of the auxiliary bearings. Support for the rotor by axiallyand radially acting auxiliary bearings during the starting and stoppingphases makes it possible to pass through oscillation-critical ranges ofspeed of rotation of the permanent magnet bearings, without involvingcriticality.

Liquid hydrogen is of very low viscosity and is a very poor lubricantand therefore imposes particular demands on the bearing arrangement offluid flow machines which operate with that medium. Due to the very lowoperating temperatures, lubrication for the bearings by means of anadditional lubricant is not possible as the lubricant would be convertedinto the solid phase. In space travel, hitherto operation has mainlybeen implemented with conventional plain or rolling bearings, the highlevel of wear of which could be tolerated in regard to the shortoperating periods and service lives which occur. That wear isproblematical in regard to planned uses for liquid hydrogen as apropellant in air travel and other areas.

The bearing arrangement in accordance with the invention for thecentrifugal or rotary pump set forth makes a contribution to increasingthe service life and operational readiness of flow machines whichoperate with liquid hydrogen.

Further advantages, features and details of the invention will beapparent from the following description of preferred embodiments andwith reference to the drawing in which:

FIG. 1 is a view in longitudinal section of a pump for conveying liquidhydrogen with a bearing arrangement according to the invention for therotor in the operative position,

FIG. 2 shows the rotor of the pump shown in FIG. 1,

FIG. 3 is a diagrammatic view showing the rotor bearing arrangementaccording to the invention in the operative position,

FIG. 4 is a diagrammatic view of another rotor bearing arrangement inthe rest position,

FIG. 5 through 7 show three diagrammatic views relating to differentpermanent magnet bearings in longitudinal section,

FIG. 8 through 10 show embodiments of the integration of a hydrodynamicplain bearing into an impeller with a cover plate of a pump,

FIG. 11 is a view in longitudinal section of a hydraulic compensatingdevice,

FIG. 12 is a view in longitudinal section of a hydrostatic plainbearing, and

FIG. 13 is a view in longitudinal section through a permanent magnetbearing with integrated front-end or intake impeller.

A centrifugal or rotary pump 10 as shown in FIG. 1 is designedespecially as an immersed pump for the delivery of liquid hydrogen. Itoperates submerged in a medium to be conveyed and in a housing 12 has aperpendicularly disposed rotor shaft 14 of a rotor 16 which is againshown separately in FIG. 2. Arranged on the rotor 16 are three half-openimpellers 18 for conveying the medium and for increasing the pressure ofthe medium.

The rotor 16 is radially supported by permanent magnet bearings 20, 22.A hydrodynamic thrust bearing 24 is disposed in the proximity of thecenter of gravity S of the rotor 16. It is a type of design for anaerodynamic plain bearing, with a supporting gap of about 1 μm. Thatprovides an adequate supporting force with the lubricant liquidhydrogen. As in the described pump the radial stiffness of the magneticbearings 20, 22 is very low with respect to the axial stiffness of thethrust bearing 24, it is possible to forego the resilient supportingarrangement, which is usual with very small supporting gaps, in respectof one of the sliding surfaces or its elements. The small supporting gapof the thrust bearing 24 and its high level of axial stiffness permitvery accurate setting of the gaps 26 between the blades of the half-openimpellers 18 and the walls of the housing. It is possible in that way tominimise the gap losses. In the operative position of the rotor 16, arotor-side bearing ring or inner magnetic ring 28 of the upper magneticbearing 22 is displaced upwardly by about 0.1 mm with respect to itsouter magnetic ring 30 as the housing-side bearing ring, so that thenecessary force for lifting the rotor 16 from the plain bearing 24 inthe stopping phase is produced. Corresponding bearing or magnetic rings32, 34 of the lower magnetic bearing 20 do not involve any displacementin the operative condition.

An auxiliary bearing 36 for limiting the axial movement of the rotor 16is also disposed in the proximity of the center of gravity S of therotor 16. It is integrated in the form of a cone-type plain bearing intothe rear of the last impeller 18 and also radially supports the rotor 16in the starting and stopping phases; radial support for the rotor 16 bythe auxiliary bearing 36 and the arrangement thereof in the proximity ofthe center of gravity of the rotor 16 make it easier to pass through thefirst rigid-body mode of the rotor 16 at lower speeds of rotation.

The rotor of an electric motor 38 which drives the pump and which is inthe form of a synchronous motor with permanent magnet excitation wasintegrated into the rotor 16 of the pump 10. A large air gap between theexciter magnet 40 and the iron return member 42 and the iron-less designof the electric motor 38 on the rotor side guarantee low unstable forcesfor the electric motor 38 in the radial direction so that radial supportfor the rotor 16 by the permanent magnet bearings 20, 22 is possible,with a relatively low level of radial stiffness. The large air gapadditionally permits the entire delivery flow to be passed through thegap, thereby guaranteeing good cooling for the electric motor 38. Themagnetic rings 28, 30, 32, 34 of the bearings 22 and 20 respectively andthe exciter magnet or magnetic ring 40 of the electric motor 38 areprotected from damage by metal protection or shroud rings 44. The latteradditionally serve to damp rotor oscillations as damping eddy currentsare induced therein upon oscillations of the rotor 16 by the fields ofthe magnets.

In order to increase the resistance to wear—and to minimise the weightof the pump 10—individual components of the rotor 16 such as impellers18 and intermediate disks or plates 45 as well as housing plates ordisks 46, 48 can be made from ceramic. The ceramic components of therotor 16 are held together by a tie bolt 50. The flexural loadingsacting on the rotor 16 during operation of the machine are carried bythe boss cross-sections of the components as compression forces.

The diagrammatic view in FIG. 3 for the bearing arrangement according tothe invention shows the rotor 16 of a fluid flow machine with impeller18 which is supported in radially contact-free manner by means of thepermanent magnet bearings 20, 22. In the operating phase of the machinethe impeller 18 generates a thrust in the axial direction, which iscarried by the hydrodynamic plain bearing 24. The components 28 of themagnetic bearing 22 on the rotor 16, upon abutment thereof against thehydraulic plain bearing 24, are arranged displaced with respect to thecomponents 30 of the magnetic bearing 22 on the housing side axiallytowards the auxiliary bearing 36 by a dimension t so that, by virtue ofthe unstable characteristic of the bearing forces of the permanentbearing magnet 22 in the axial direction, there is a magnetic force A inthe axial direction, which is in opposite relationship to the thrustforce B of the impeller 18. When the rotor 16 is lifted off the plainbearing 24, at the permanent magnet bearing 20 there is also an axialdisplacement of the components 32 on the rotor side with respect to thecomponents 34 on the housing side. The resulting forces in the axialdirection of the magnetic bearings 20, 22 increase with an increasingspacing of the rotor 16 from the plain bearing 24.

The axial displacement of the bearing components is limited by theauxiliary bearing 36 which in FIG. 3 is in the form of a cone-type plainbearing. The rotor 16 of the fluid flow machine is freely movablebetween the conditions of abutment against the bearings 24 and 36. Thearrangement of the bearing elements of the bearings 20, 22, 24, 36 issuch that the axial force generated by the permanent magnet bearing 22is sufficiently great to lift the rotor 16 off the hydrodynamic plainbearing 24 in the stopping phase of the machine. The displacement of therotor 16 is limited by the auxiliary bearing 36 so that the minimumaxial thrust generated by the impeller 18 in the operative phase isgreater than the maximum force generated in the axial direction by thepermanent magnet bearings 20 and 22 and which occurs when the rotor 16abuts against the auxiliary bearing 36. That ensures that in operationthe rotor 16 lifts off the auxiliary bearing 36 and is supported axiallyby the hydrodynamic plain bearing 24 without contact between the slidingsurfaces. In the stopping phase the rotor 16 is lifted off thehydrodynamic bearing 24 again by the magnetic forces and is pressedagainst the auxiliary bearing 36. A suitable choice in respect of thebearing materials and a small diameter for the auxiliary bearing 36ensure a low starting torque for the bearing arrangement.

In the embodiment shown in FIG. 4 the rotor 16 is in the rest positionand the rotor-side inner magnetic ring 28 of the magnetic bearing 22projects on the impeller side by the dimension t. In this case, themagnetic force A is in a direction towards the impeller 18. Here, themagnetic force A urges the rotor 16 out of the auxiliary bearing 36 intothe operative position. The drawing does not show that a lifting devicesuch as an electromagnetic unit is required for lifting the arrangementout of the operative position.

FIG. 5 diagrammatically shows on a rotor axis or shaft 14 a magnetic orbearing ring 28 near to the rotor, with an outer magnetic or bearingring 30 which surrounds same and which is axially displaced by thatdimension t; therein the magnetisation direction is indicated by arrows52. This bearing arrangement is based on the repulsion forces of themagnets 28, 30.

In the case of bearings as shown in FIG. 6 which operate with shearingforces, the magnetic rings 28, 30 on the rotor and housing sides aregenerally disposed at the same height, that is to say not inwardly oroutwardly, that is to say at different spacings relative to thelongitudinal axis Q.

In the case of so-called reluctance bearings as shown in FIG. 7, thesupporting action can also be achieved if magnets are arranged only onthe housing side or only on the rotor side. In that case, onlyferromagnetic material—generally iron 54—is provided on the oppositeside. The fixing to the housing 12 is indicated in broken line at 13.

In the drawing the dimension t of the displacement is determined by themagnetic rings 28, 30 of a magnetic bearing. In FIG. 6 the dimension tis defined by way of the two spacings t1 and t2 which can be seentherein: t=(t1−t2)/2.

FIGS. 8 through 10 diagrammatically show the integration of ahydrodynamic plain bearing 24 into an impeller 18 with a cover plate ordisk. A major advantage of integrating the plain bearing 24 into animpeller 18 is its particular sealing action—similarly to that of asliding ring seal with hydrodynamically repelling sliding surfaces—andthus the suitability of the plain bearing 24 as a slit-type or labyrinthseal for the impeller 18.

FIG. 8 shows the integration of a hydrodynamic plain bearing 24 whichhas a purely axial effect, into a closed impeller 18. When aerodynamicplain bearings are integrated into a plurality of impellers 18 of arotor 16. all plain bearings, except for one, must be constructed withresiliently supported sliding or bearing surfaces. Suspending the rotor16 with a low degree of stiffness by virtue of the permanent magnetbearings 20, 22 can be utilised only for an individual plain bearing, tocompensate for positional errors in respect of the sliding surfaces. InFIG. 9 the sliding surface on the housing side and in FIG. 10 thesliding surface on the impeller side of the plain bearing 24—in thiscase the entire impeller 18—is shown as being resiliently supported.

FIGS. 9 and 10 each show at 56 an elastic ring, while in addition thehydrodynamic bearings 24 in FIGS. 9 and 10 are in the form of cone ortaper bearings. In that way it is additionally possible to achievehighly precise radial guidance for the rotor 16 in operation as well asprecise axial guidance.

FIGS. 11 and 12 each diagrammatically show an example of a hydrauliccompensating device—compensating plate or disk 58 (FIGS. 11) —and for ahydrostatic plain bearing 60. Reference 62 in FIG. 11 denotes an annulargap while reference 64 in FIG. 12 denotes a bearing disk with pressurefeed 66.

In FIG. 11 the compensating disk 58 is subjected in operation on theimpeller side to the action of the delivery pressure, while the otherside is subjected to the action of the ambient pressure or a lower stagepressure. The annular gap 62 is of such a size that in the operativephase a movement of the impeller 18 due to the axial thrust in thedirection of the suction or intake side results in a reduction in thegap width and the resulting increase in pressure in the gap provides forcompensation in respect of the axial thrust, without the gap widthfalling below a minimum value.

The hydrostatic plain bearing 60 in FIG. 12 can be disposed for exampleon the impeller 18 of the first stage. The supply of pressure for theplain bearing 60 can be implemented in the operative phase by returningthe stage pressure from a higher stage.

FIG. 13 shows the integration of a front-end or intake impeller 68 intothe interior of a magnetic bearing 28, 30. This arrangement saves onspace, it optimises the feed of the medium being conveyed (axial feed:improved suction intake capability), while the impeller 68 is lesssensitive to radial displacement of the rotor 16. The same is alsopossible for a normal axial impeller.

What is claimed is:
 1. A fluid flow apparatus, comprising: a fluid flowmachine assembly including a rotor and at least one impeller which, inan operating condition, generates an axially directed thrust; at leastone hydrodynamic plain bearing for carrying said axially directed thrustand for axially guiding said rotor; and at least one permanent magnetbearing for radially guiding said rotor, said at least one permanentmagnet bearing having a rotor component and a fixed component, saidrotor component being displaced axially with respect to said fixedcomponent so as to produce a force opposite to said axially directedthrust.
 2. An apparatus as set forth in claim 1, wherein slidingsurfaces of said at least one hydrodynamic plain bearing which areoperative in said operating condition are separated from each otherduring a stopped condition and during starting and stopping phasesbetween said operating condition and said stopped condition, and furthercomprising at least one additional bearing which guides said rotorduring said starting and stopping phases, said at least one additionalbearing being adapted so that in said stopped condition, said at leastone additional bearing has a lower moment of friction than said at leastone hydrodynamic plain bearing.
 3. An apparatus as set forth in claim 2,wherein said at least one additional bearing and said at least onehydrodynamic plain bearing define an extent of axial displacement ofsaid rotor, and wherein said force of said at least one permanent magnetbearing is less than a minimum axially directed thrust of said fluidflow machine assembly in said operating condition.
 4. An apparatus asset forth in claim 1, wherein said rotor component is a rotor-sidebearing ring, and wherein said fixed component is a housing-side bearingring.
 5. An apparatus as set forth in claim 1, wherein said fluid flowassembly further includes a motor defining a motor gap, and wherein aflow path for fluid driven by said fluid flow assembly passes throughsaid motor gap.
 6. An apparatus as set forth in claim 1, wherein saidrotor component and said fixed component define a bearing gap, andwherein a flow path for fluid driven by said assembly passes throughsaid bearing gap.
 7. An apparatus as set forth in claim 1, wherein saidat least one hydrodynamic plain bearing is an aerodynamic plain bearing.8. An apparatus as set forth in claim 1, wherein said hydrodynamic plainbearing is connected to a cover disk of an impeller of said fluid flowmachine assembly.
 9. An apparatus as set forth in claim 8, wherein saidat least one hydrodynamic plain bearing has a rotor-side sliding surfaceand a plain bearing, wherein said rotor side sliding surface isintegrated into said impeller, and wherein said rotor-side slidingsurface and said plain bearing are sealing with respect to a highpressure side and a low pressure side of said impeller.
 10. An apparatusas set forth in claim 1, wherein said rotor is radially supported bysaid at least one permanent magnet bearing, and wherein saidhydrodynamic thrust bearing is positioned substantially adjacent to acenter of gravity of said rotor.
 11. An apparatus as set forth in claim10, further comprising an auxiliary bearing provided substantiallyadjacent to said center of gravity of said rotor, wherein said auxiliarybearing is a cone plain bearing.
 12. An apparatus as set forth in claim11, wherein said auxiliary bearing is integrated into a rear surface ofsaid impeller.
 13. An apparatus as set forth in claim 1, wherein said atleast one hydrodynamic plain bearing is an aerodynamic plain bearinghaving a supporting gap of about 1 μm.
 14. An apparatus as set forth inclaim 1, further comprising an electric motor for driving said fluidflow machine assembly, said electric motor having a motor rotor which isintegrated into said rotor of said fluid flow machine assembly.
 15. Anapparatus as set forth in claim 14, wherein said electric motor is asynchronous motor having permanent magnet excitation.
 16. An apparatusas set forth in claim 15, further comprising protective rings forprotecting bearing rings of said at least one permanent magnet bearingsand magnetic rings of said electric motor.
 17. An apparatus as set forthin claim 16, wherein said protective rings are provided from a materialselected from the group consisting of carbon fiber-reinforced plasticmaterial, glass fiber-reinforced plastic material and combinationsthereof.
 18. An apparatus as set forth in claim 1, wherein at least oneof said rotor and said impeller are made from ceramic material.
 19. Anapparatus as set forth in claim 1, wherein said at least onehydrodynamic plain bearing comprises bearing surfaces which, in saidoperating condition, axially guide said rotor without contact betweensaid bearing surfaces.
 20. An apparatus as set forth in claim 1, furthercomprising at least one of a hydrostatic plain bearing and a hydrauliccompensating device supportingly associated with said at least onehydrodynamic plain bearing.